Compression intercooled gas turbine combined cycle

ABSTRACT

A compression intercooled gas turbine and vapor bottoming combined cycle system with the gas turbine operating at 30 to 65 atmospheres is disclosed. A twin spool hot gas generator incorporates compression intercooling at the optimum intercooler pressure ratio to (a) minimize intercooler heat rejection degradation, (b) raise the overall cycle pressure ratio, (c) increase gas generator core mass flow and (d) to increase the gas turbine power output. The gas turbine can operate in either the simple cycle or the reheat cycle mode for optimum combined cycle efficiency. A combined cycle efficiency upwards of 60 percent can be realized.

CROSS REFERENCE TO RELATED APPLICATIONS

This application is a continuation of application Ser. No. 818,472 filedJan. 13, 1986, (abandoned), which is a continuation-in-part of U.S.application Ser. No. 047,571, filed June 11, 1979, now U.S. Pat. No.4,314,442.

This application is also a continuation-in-part of U.S. application Ser.No. 224,496, filed Jan. 13, 1981 (U.S. Pat. No. 4,438,625), which is adivision of U.S. application Ser. No. 954,832, filed Oct. 26, 1978, nowU. S. Pat. No. 4,272,953.

This application is also a continuation-in-part of U.S. application Ser.No. 274,660, filed June 17, 1981, now U.S. Pat. No. 4,384,452, which isa division of U.S. application Ser. No. 047,571, filed June 11, 1979,now U.S. Pat. No. 4,314,442.

This application is also a continuation-in-part of the four U.S.application Ser. Nos. 416,171 (abandoned); 416,172 (abandoned); 416,173(U.S. Pat. No. 4,550,562) and 416,275 (U.S. Pat. No. 4,507,914), filedSept. 9, 1982, with each of these being a continuation-in-part of theheretofore above stated U.S. Applications and U.S. Patents.

This application is likewise a Continuation-in-Part of the three U.S.application Ser. Nos. 486,334 (U.S. Pat. 4,545,197); 486,336 (U.S. Pat.No. 4,565,490) and 486,495 (U.S. Pat. No. 4,543,781) filed Apr. 19,1983, with each of these being a continuation-in-part of the heretoforeabove stated U.S. Applications and U.S. Patents.

BACKGROUND OF THE INVENTION

This invention relates to an open-cycle compression-intercooled gasgenerator operating at high cycle pressure ratios of 35 to 65atmospheres. The said gas generator exhausts hot gas at relatively highvelocity and pressure to a diffuser where a majority of the velocitypressure is converted to static pressure. The converted gas then isreheated in a reheat combustor before being expanded through a powerturbine to produce mechanical work, generally considered to beelectrical power.

The air is intercooled at a particular and specific pressure to minimizethe overall combined cycle efficiency degradation when said gasgenerator, diffuser, reheat combustor and power turbine are operating inconjunction with a heat recovery boiler and a steam or vapor turbine.The boiler can evaporate water, ammonia, freon or some other liquid or amixture thereof to form superheated vapor for expansion through a vaporturbine such as a steam turbine. A combined cycle is accordingly formedwhereby the overall combined cycle efficiency can range from 50 to 65%(LHV) depending upon the gas generator and power turbine inlettemperatures and the bottoming steam or vapor cycle selected.

An integral single-bodied gas generator with a coaxial shaftingarrangement for driving the low pressure and high-pressure compressorsections with intercooling connections is made possible whereby gasgenerators such as the GE LM5000, RR RB211, P&WA JT9 and subsequentthird-generation gas generators expected to be developed from the NASAE³ (Energy Efficient Engine) aircraft gas turbine program can be appliedthrough proper modifications. The advantage of quick installation andremoval of aeroderivative gas generators can be retained.

A pressure retaining casing arrangement around the high pressure portionof said gas generator, that is the high pressure compressor, thecombustor and the turbine section, is provided to contain the higherthan normal pressure of said gas generator. The GE LM5000 operates at acycle pressure ratio of about 30 and the new engines to be derived fromthe E³ program will operate at pressure ratios of about 38 atmospheres.The gas generator of this invention would operate at pressures of 35 to65 atmospheres, but preferably at about 50. The pressure-retainingcasing makes it possible to adapt light-weight aero-derivative gasgenerators with light-weight casings for the higher pressure levels ofthis invention without exceeding safe blow-out casing pressures.

The reheat-gas-turbine combined cycle is being seriously considered as away to obtain a higher combined-cycle efficiency than otherwiseobtainable from the simple-cycle gas turbine. The Japanese Government iswell along in testing its 122 MW compression-intercooled reheat gasturbine and field-test results will be made available in mid-1984.

The Japanese reheat-gas-turbine configuration incorporates compressionintercooling to accomplish a projected 55 percent (LHV) combined-cycleefficiency. Intercooling is done at about a 4.85 ratio by direct-contactcondensate spray water. The intercooling of my invention takes place ata much lower and specific pressure with an optimum ratio range of about2.0 to 2.5 and uses condensate and cooling-tower, lake, river or seawater as the coolant in a primary closed loop and an open or closedsecondary loop. Ammonia or a mixture of ammonia and water can also beused as the intercooler coolant where a dry-type atmospheric coolingsystem is employed. Direct-contact water-spray cooling can also be used.

Studies of the non-intercooled reheat-gas-turbine combined cycle showthat such an arrangement will produce the highest combined-cycleefficiency for any given gas generator and reheat-turbine inlettemperatures. However, the nonintercooled gas generator is limited toabout 40 atmospheres primarily due to the high compressor dischargetemperatures associated with the high compressor pressure. My inventionmakes it possible to exceed the 40 atmospheres and provide a lowercompressor-discharge temperature needed for said gas generatorcombustor-liner cooling and NO_(x) control.

Industrializing the E³ engine gas generators, considered to be the thirdgeneration of aircraft engines, is inevitable based on past history, andadapting them for intercooling can be accomplished using the basicengine designs coming from the E³ program and applying the processprinciples and design features of this invention. Adaptation of the E³engine such as the Pratt and Whitney Aircraft 2037 and 4000 seriesengines as well as similar engines by General Electric and Rolls Roycebeing readied for aircraft service can be applied.

In U.S. Pat. No. 4,272,953 applicant has disclosed that secondgeneration, high-cycle pressure ratio, high-firing temperature gasgenerators can be used in the reheat gas turbine/steam turbine combinedcycle to yield increased efficiency and output heretofore unexpectedfrom reheat-gas-turbine combined cycles. A novel reheat gas turbinewithout intercooling combined with a steam turbine is further disclosedin applicant's pending application, U.S. Ser. No. 224,496 filed Jan. 13,1981. In this pending application the reheat gas turbine comprises ajuxtaposed and axially aligned gas generator and power turbine in whichgas flow through the gas generator, reheat combustor and power turbineis substantially linear throughout, but nothing is given on thecompression intercooling in either disclosure.

Other U.S. patents and pending applications by the applicant, allpertaining to the reheat gas turbine and steam cooling, but notspecifically to compression intercooling are as follows:

U.S. Pat. No. 4,314,442

U.S. Pat. No. 4,384,452

U.S. Ser. No. 416,171

U.S. Ser. No. 416,172

U.S. Ser. No. 416,173

U.S. Ser. No. 416,275

U.S. Ser. No. 486,334

U.S. Ser. No. 486,336

U.S. Ser. No. 486,495.

Intercooling has been used for many years with compression of air andother gaseous fluids to reduce the power required for compression. Alsosimple-cycle and reheat-cycle gas turbines have incorporated aircompression intercooling to reduce compression work and consequently toincrease the gas turbine output, particularly for regenerative cycle gasturbines not involving a combined cycle. However, in such cases theemphasis has been on maximizing output and gas turbine efficiency andnot to optimize combined cycle efficiency. As will be shown, thecompression intercooling in past gas turbines takes place at a muchhigher pressure ratio than the analytical discovery of my inventionindicates as being optimum for combined cycle efficiency. The JapaneseGovernment is developing an intercooled reheat gas turbine for combinedcycle service, but water-spray intercooling is employed at a much higherintercooled compression ratio than that of my invention.

A coaxial shafting arrangement is contemplated whereby the initial (lowpressure) compressor is driven by a turbine by means of a shaft runningthrough the high-pressure compressor, the high-pressure turbine and theinterconnecting shaft. Coaxial drives are highly developed forhigh-bypass fan jets and indeed gas generators such as GeneralElectric's LM5000, General Motors 570K and Rolls-Royce's RB-211.However, no intercooling is used or even remotely suggested. It is theincrease in the low-pressure compressor diameter made possible by theintercooling that permits adequate physical space for the air to beexited and readmitted efficiently with a minimum of pressure loss.

Further, when such type generators as the LM5000 or 570K or futurethird-generation aeroderivatives are modified for intercooling, thehigh-pressure sections (high-pressure compressor housing, combustorhousing and high-pressure turbine housing) are subjected to much higherinternal pressures. The cycle pressure ratios will increase from 18, 30or 38, as the case may be, to some 35 to 65 due to the supercharger. Inorder for such gas generators to be adapted for compression intercoolingand the higher cycle pressure ratios suitable for the reheat gas turbineand the combined cycle something has to be done about the added pressureto prevent casing rupture and/or distortion due to the higher internalpressure. This invention deals with the added pressure and theprevention of distortion and blow-out by incorporating a special andunique cylindrical pressure chamber with thermal expansion joints aroundthe gas generator. Air is presently being used to cool the advanced aerogas-turbine casings to control rotating blade-tip clearance. Thisinvention uses steam inside the pressure chamber not only to providecasing cooling but to also provide cooling for the internalgas-generator parts.

SUMMARY OF THE INVENTION

This invention contemplates a process and apparatus for generating ahigh-pressure, high-temperature gas to be reheated and expanded in apower turbine whereby useful mechanical work is produced. Additionalpower is also produced by a steam or vapor turbine operating in acombined cycle mode using the heat from the gas-turbine exhaust togenerate the steam or vapor.

The process comprises the compression of air in a low-pressurecompressor at a specified pressure ratio, intercooling said air by heatexchange or direct contact and then compressing said air to a higherthan normal pressure whereafter said compressed air is heated directlyby fuel and partially expanded in a gas generator arrangement. Thegas-generator exit gas is further reheated by fuel and further expandedthrough a power turbine to produce mechanical work before beingexhausted to a heat recovery boiler. Steam or vapor generated in saidheat recovery boiler drives a steam or vapor turbine to producesecondary mechanical work.

Accordingly, it is an object of this invention to provide an intercooledgas generator of a higher than normal cycle pressure ratio of 35 to 65atmospheres, preferably about 50, with a coaxial shafting arrangement asan integral modular unit for easy installation and removal with respectto the intercooler(s), gas-generator exit diffuser, reheat combustor andpower turbine.

A further object is to provide a gas generator with a preferred cyclepressure ratio of about 50 atmospheres for a gas-generator firingtemperature of up to 2600° F. or even higher.

A further object of this invention is to provide an intercooled gasgenerator with a much larger than normal low-pressure compressorpitch-line diameter to allow adequate space for an exit elbow and radialdiffuser and an associated low-pressure-drop return ducting to thehigh-pressure compressor of said gas-generator module.

Still another object of the invention is to provide acylindrical-pressure container arrangement around the high-pressuresection of the gas-generator module to allow steam pressure and/or bleedair pressure to pressurize the outer gas-generator casing thus allowinglight-weight aero-derivative type gas generators to be adapted for muchhigher than normal pressure ratios without danger of pressure rupture orcasing distortion. Thermal expansion joints are provided in the pressurecontainer to compensate for differential thermal expansion.

A further object of the invention is to provide an annular plenum aroundthe high-pressure section of said gas-generator module whereby coolingsteam can be admitted to the interior of said gas generator to cool thestationary parts and nozzle vanes and rotating disks and blades wherebymultiple separate inlet headers and connectors are eliminated.

Still another object of this invention is to provide a process forintercooling the air at the proper pressure ratio to optimizecombined-cycle efficiency and minimize combined-cycle efficiencydegradation. Heat exchange or direct-contact waterspray cooling iscontemplated.

A further object of the invention is to provide alow-pressure-compressor cascade-airfoil diffuser and an associatedprocess to recover additional dynamic pressure otherwise lost to thecycle. An incrementally higher combined-cycle efficiency is obtained.

Another further object of the invention is to provide gas-generatorouter-casing cooling to control thermal growth of the casing and tocontrol rotating blade-tip clearance.

Also contemplated within the scope of the invention is a combinedintercooled reheat gas turbine and steam or vapor turbine cycle forproduction of useful power wherein superheated steam or vapor isproduced by heat exchange in either the powerturbine exhaust-ductsection or the reheat combustor wherein said superheated steam or vapordrives a steam or vapor power turbine for production of additionaluseful power to that of the power produced by the reheat gas powerturbine. The steam or vapor can optionally be extracted, reheated andreadmitted to the steam or vapor turbine and normally condensed forhigher cycle efficiency and increased incremental power output, whichwill become subsequently apparent reside in the details of constructionand operation as more fully hereinafter described and claimed, referencebeing had to the accompanying drawings forming a part hereof, whereinlike numerals refer to like parts throughout.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic view of a compression intercooled gas generatorexhausting through a diffuser to a reheat combustor and then to a powerturbine to drive a mechanical or electrical load with exhaust gassesbeing ducted to a heat recovery means for forming a combined cycle.

FIG. 2 is a top plan view, partially in diagramatic section, of theintercooled gas generator of the present invention illustrating theintercooling exit and re-entrance arrangement.

FIG. 3 is an enlarged top plan view in partial section of FIG. 2 showingmore details.

FIG. 4 is a top plan view in partial section showing increase in thelow-pressure compressor pitch-line radius resulting from intercooling.

FIG. 5 is an enlarged plan view in section of FIG. 2 showingpressure-casing expansion joint and cooling arrangement.

FIG. 6a is a perspective view of the gas generator high-pressure sectionshowing low-pressure compressor on the left and diffuser, reheatcombustor and power turbine on the right.

FIG. 6b is a perspective view with partial section ofcompression-intercooled reheat gas turbine incorporating an industriallow-pressure compressor in the gas generator.

FIG. 7a is a pan quarter top sectional view showing an airfoil diffuserand toroidal-shaped outlet duct.

FIG. 7b is an enlarged sectional view of the diffuser cascade airfoilsof FIG. 7a.

FIG. 8 is a front quarter top sectional view of FIG. 7a viewed indirection of axial air flow.

FIG. 9 is a graph of combined-cycle efficiency as a function ofcycle-pressure ratio, firing temperatures and steam conditions.

FIG. 10 is a graph showing the relationship of compressor temperaturesto cycle-pressure ratio.

FIG. 11 is a graph showing incremental cycle efficiency versusintercooler pressure ratio for a 40 cycle pressure ratio condition.

FIG. 12 is a graph showing incremental cycle efficiency versusintercooler pressure ratio for a 60 cycle pressure ratio condition.

FIG. 13 is a graph showing incremental work saved by intercooling as afunction of cycle pressure ratio for 40 and 60 cycle pressure ratioconditions.

FIG. 14 is a graph showing percent change in combined cycle efficiencyas a function of intercooler pressure ratio for 40 and 60 cycle-pressureratio conditions.

FIG. 15 is a graph showing gas generator combustor inlet temperaturerise and compressor discharge temperature as a function of intercoolerpressure ratio for 40 and 60 cycle-pressure ratio conditions and 2600°F. gas inlet temperature.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

In FIG. 1 a schematic diagram of the intercooled gas generator of thepresent invention shows gas generator 20 which receives air throughinlet line 22 producing compressed air by low-pressure compressor 24which is driven through A shaft 26 by low-pressure gas-generator turbine28 which is powered by heated gas produced in first combustor 30 fromair entering combustor 30 through line 32 and fuel entering combustor 30through fuel line 34. Low-pressure air is discharged through line 36 andis cooled by series intercoolers 38 and 40 before being dischargedthrough line 42 to the high-pressure compressor 44. The air is furthercompressed by high-pressure compressor 44 driven by high-pressureturbine 46 through B shaft 48. High-pressure turbine 46 is powered bysaid heated gas produced in the first combustor 30. Reheat or secondcombustor 50 receives exhaust from gas-generator turbine 28 throughreheat diffuser 52 and fuel through line 54 and discharges reheated gasthrough line 56 to power turbine 58 which drives load 60, preferably anelectric generator, directly by C shaft 62. The gas exits power turbine58 through line 63 to a heat-recovery boiler where steam or a vapor suchas freon or a mixture of water and ammonia is produced to drive a secondsteam or vapor turbine, forming what is commonly referred to in theindustry as a combined cycle. Refer to FIG. 1.

Power turbine 58 discharges hot exhaust gasses through line 63 to heatrecovery means 49 and exits said gas through line 71. Steam or vaporgenerated in heat recovery means 49 is expanded through turbine 55 whichdrives load 57. Expanded steam or vapor is condensed in condenser means61. Optionally, condensate is pumped through intercooler means section38 to cool compressed air entering said intercooler means through line36. Heated condensate returns to heat recovery means through line 69.Lines 73 and 75 feed steam to gas generator 20 and power turbine 58.

Low-pressure compressor 24 discharge air can be cooled first bycombined-cycle steam or vapor condensate in a counterflow direction withthe coolant entering intercooler 38 through line 64 and dischargingthrough line 66. The air can be further cooled by cooling-tower, lake,river or sea water or a coolant from a dry-cooling system with thecoolant counterflowing and entering second series intercooler 40 throughline 68 and discharging through line 70. Direct-contact water-spraycooling can also be applied.

It is to be particularly noted that the component parts of coaxial-shaftgas generator 20 shown in FIG. 1 are conventional and only theparticular physical modular arrangement and method of getting thepartially compressed air out of and back into said coaxialshafted-modular gas-generator compressor 24 and 44 at a specificlow-pressure range and the physical arrangement and method ofcontrolling blow-out pressure and the introduction of coolant to saidmodular gas-generator internal and external parts as well as theinterrelated process involved with the combined cycle leading toadvantages and efficiencies disclosed in the present invention areintended to be described as new.

Fuel for combustors 30 and 50 can be liquid such as petroleumdistillates, crude oil, Bunker "C" or petroleum liquid products such asmethanol,or said fuel can be gaseous such as natural gas or gas producedfrom coal (low, medium or high BTU) and said fuel can be burned inconjunction with an integrated coal gasification combined gasturbine/steam turbine power plant.

FIGS. 2 and 3 are overall representations of the typical modular gasgenerator 20 showing inlet line 22, low-pressure compressor 24 andassociated outer casing 72, variable stator blades 74,variable-stator-blade linkage mechanism 76, discharge duct 36, returnduct 42, high-pressure compressor 44, associated outer casing 78,combustor 30, high-pressure turbine 46, low-pressure turbine 28, coaxialshafting 26 and 48 and exit 51 to diffuser 52.

There can be two discharge ducts 36, one on each side of gas generator20 and two return ducts 42, one on each side of gas generator 20 to feedtwo mirror-image twin intercoolers 38 and 40 located on each side of gasgenerator 20. The dual ducting and intercooler arrangement makes itpossible to manage the flow of the low-density high-volume air to andfrom the intercooler(s) more efficiently with a lower overall pressureloss.

Referring to FIG. 3, an enlarged partial view of FIG. 2, thelow-pressure compressor 24 is supported by forward antifriction steelball bearing 80 and is cantilevered or overhung. Said compressor rotordrum 114 is mounted to shaft 26 by flange 82. The high-pressurecompressor 44 is supported by duplex antifriction bearing(s) 84 forcoaxial shafts 26 and 48. The high-pressure turbine 46 and low-pressureturbine 28 are supported by rear duplex antifriction bearing(s) 86. Thelow-pressure shaft 26 and compressor 24, components for a gas generatorwith an airflow of 300 lb/sec, would typically rotate at about 3600 RPMand the high-pressure shaft 48 and compressor 44 would typically rotateat about 8500 RPM. Note that hydrodynamic-type bearings (usually babbit)can be applied for industrial-type construction.

FIG. 3 is an enlarged view of FIG. 2 representative of the typicalhigh-cycle pressure-ratio intercooled gas generator 20 of the presentinvention including low-pressure compressor 24, high-pressure compressor44, low-pressure turbine 28, high-pressure turbine 46, coaxial shaftingA and B (26 and 48), outlet ducting 36, return inlet ducting 42 and exit51 to diffuser 52. Gas-generator low-pressure compressor 24 is made upof stages 88, 90, 92, 94, 96, 98 of a six-stage axial-flow compressor.There can be fewer stages or more stages as required to produce thedesired pressure ratio, and for a typical pressure ratio of 2.5 therewould be four to six stages according to the particular design andmanufacturer.

Low-pressure compressor 24 discharges axially at about Mach 0.3 airvelocity corresponding to about 400 to 450 ft/sec at the flowingtemperature. Said air flows into elbow 100 where the airflow directionis changed from that of axial to radial. The air flow at this point isthen diffused by flat or parallel wall vaneless diffuser 102 beforebeing discharged to a scroll collector 35 and then to duct 36. The gasvelocity exiting diffuser 102 is at about Mach 0.14 corresponding toabout 200 ft/sec according to the base temperature involved. About 65 to70 percent of the velocity pressure head is recovered by diffuser 102and the remainder is dissipated and lost to the cycle. The cooled airenters the high-pressure compressor 44 through a separate duct 42 bymeans of a special "S" shaped inlet duct 104 detachably attached tocompressor outer casing 78, inner housing 79 and elbow 100. Struts 77joins together forward and aft sides of inlet duct 104. The cooled aircounterflows in direction to that of the hot discharge air flowinginside diffuser 102. Diffuser 102 and elbow 100 fit adjacent and closeto the outer wall 108 of inlet duct 104 and are removably securedtogether. Generous cross-sectional area of the inlet duct 104 isprovided to keep the inlet air velocity low at a value of approximately50 to 75 ft/sec to reduce inlet pressure loss.

The added low-pressure compressor 24 pitch-line radius, r, FIG. 2, madepossible by the cooler and more dense air to the high-pressurecompressor 44 provides the needed radial space and therefore makes itpossible to exit the low-pressure compressor 24 hot air, diffuse the hotair and return the cooled air to the high pressure compressor 44efficiently and with low-pressure loss and at the same time retain thecoaxial shafting and unitized single-unit modular construction of gasgenerator 20. It is obvious that if the pitch-line radius, r, of thelow-pressure compressor 24 were the same or very nearly the same as thepitch-line radius of the high-pressure compressor 44, there would not beadequate radial space without the shafting 26 to the low-pressurecompressor being lengthened considerably making it impractical from astructural standpoint to adapt an aero-derivitating gas generator to aunitized single module and structural design with the given bearingarrangement provided.

Reference is made to FIG. 4 which shows the low pressure compressor 24of this invention with radius r₂ superimposed over the low-pressurecompressor or a normal gas generator with radius r₁. Also note thesmaller pitch-line radius, r₃, of the high-pressure compressor 44. Aspreviously mentioned, the greater low-pressure compressor 24 radius ismade possible by the change in density of the air when it isintercooled. The forward-bearing arrangement is basically the same forboth cases and the stationary structural support to the bearing needsonly minor changes to accomodate the inlet "S" shaped duct 104 whereinthe duct 104 becomes part of the stationary structural part of saidmodular gas generator 20. The low-pressure compressor supportfunnel-shaped rotating members 110 and 112 likewise change only slightlyin shape to accomodate the inlet duct 104 and low-pressure-compressorrotor drum 114. Note that the low-pressure compressor drum 114 ismounted to shaft 26 similar to how a dual truck tire and rim arecantilevered and bolted to an axle shaft which protrudes beyond thewheel-axle bearing.

The thermodynamic explanation of the increase in radius, r, will bediscussed as follows: The high-pressure compressor 44 running at aconstant RPM has an inlet volume flow that remains practically constant,if not constant, for varying conditions of inlet pressure andtemperature, that is changes in its inlet density. The air that iscompressed by the low-pressure compressor 24 is cooled before it isadmitted to the inlet of the high-pressure compressor 44 and as a resultthe density is increased according to Boyles Law given as follows:

    PV=WRT                                                     (1)

where P is the absolute pressure, V is the volume, W is the weight flow,R is the gas constant and T is the absolute temperature.

As an example, if the low-pressure compressor 24 discharge air is cooledfrom 300° F. to 80° F. then the change in density, according to formula(1) would be 41 percent. The low-pressure compressor 24 would have tocompress 41 percent greater mass of air to satisfy the high-pressurecompressor 44 inlet volume. Considering no increase in low-pressurecompressor 24 RPM then the pitch-line radius, r₁, of a normallow-pressure compressor would increase by about 19 percent to r₂ (squareroot of 1.41) See FIG. 4. Dynamic similarity principles apply.

There is another significant difference which can increase thelow-pressure compressor 24 pitch-line radius, r, even further. Asanother example, in the case of the GE LM5000 gas generator converted tointercooling, the front low-pressure compressor would normally have a1.667 pressure ratio and a 152° F. discharge temperature to thehigh-pressure compressor 44 for a 59° F. inlet temperature. Thehigh-pressure compressor 44 has an 18 pressure ratio and discharges at30 atmospheres for a 30 total ratio. If the low-pressure compressor ismodified to discharge at 2.5 atmospheres to form low-pressure compressor24 instead of its normal 1.667, then the discharge temperature would beabout 236° F. for a 59° F. ambient inlet. The inlet mass flow to the"core" of high-pressure compressor 44 would be 1.59 times as great ifthe compressed air is cooled to 100° F., considering a 3 percentpressure drop in the cooling process. The pitch-line radius, r₂, FIG. 4,would increase by about 1.26 times.

It is of particular interest to note that the change in radius of r₁ tor₂ with respect to high-pressure compressor radius 44, r₃, will be about75 percent; that is considering r₁ to be unity in length, r₂ would beapproximately 2 units in length and r₃ would be approximately 23/4 unitsin length or 3/4 of a unit longer than the one unit length of r₃ as canbe seen in FIG. 4. The added length of radius r₂ is critical in makingit possible and practical to retain one single module of gas generator20 and to provide the outlet elbow 100, diffuser 102 and the return "S"shaped inlet duct 104 all with the shafting 26 and 48 and forwardbearing 80 arrangement confinements of the aero-derivative design. Theradial length of the low-pressure compressor 24 radius r₂ can, however,range from 2.2 to 3.0 times that of the high-pressure compressor 44radius r₃ to provide room for said inlet and exit ducting.

The "core" gas-turbine portion of the gas-generator module, that is theintegral assembly consisting of the high-pressure compressor 44, thecombustor 30, shaft 48, the high-pressure turbine 46 and associatedcasing can have a pressure ratio of 12 to 24 according to its designwith the preferred ratio range being 18 to 24. The "core" portion canalso have two spools such as the Rolls-Royce RB-211 with twocompressors, turbines and shafts to form the high-pressure section andthus the "core". The LM5000 "core" (one-spool design) has an 18compression ratio and the new E³ "core" engine (also a one-spool design)has a compression ratio of about 23, giving examples. The low-pressurecompressor 24 of this invention would have a compression ratio of 1.8 to3.2 to produce a 50 total pressure ratio, this total ratio consideredpreferable for this invention although a total ratio as low as 35 or ashigh as 65 can also be considered and accomplished by varying thecompression ratio of the low pressure compressor 24, the total ratiobeing the product of the high and low pressure ratios. It should benoted that the low-pressure compressor 24 and associated low-pressureturbine 28 are not extremely high-technology, high-temperature parts,and these parts can be readily modified by adding stages and diameter toobtain the required pressure ratio.

The mass-flow increase to the "core" portion of the gas-generator moduleis significant because the high-technology "core" portion would be a keypart of a new gas generator 20 system developing significantly morepower output in terms of discharge pressure and mass flow to the reheatcombustor and power turbine 58, this power being in all practicalpurposes directly proportional to the increase in "core" massflow. Therewould also be an additional increase in gas generator power outputbecause of the power saved by the intercooled compression process whichwill be discussed later.

The partially compressed and cooled air enters the high-pressurecompressor 44 by way of inlet duct 104 and is further compressed byaxial-flow compressor stages 116, 118, 120 and subsequent stages of atypical eleven-stage compressor to a high-pressure compression ratio of12 to 24 as the case may be and as required. Additional high-pressurecompressor stages can be added as required. The high-pressure compressor44 can be equipped with variable stator blades 122 and mechanism 124 toprovide operating flexibility.

The high-pressure compressed air flows through gas-generator combustor30 and is heated by fuel from fuel lines 34 before being expandedthrough the high-pressure turbine 46 and low-pressure turbine 28. FIG. 3shows typically one turbine stage for the high-pressure nozzle vanes126, rotating blades 128 and disk 130 but it should be noted that two ormore stages can be applied. Similarly, one turbine stage of low-pressurenozzle vanes 132, rotating blades 134 and disk 136 are typically shownand two or more stages can be utilized. The hot gas exits by duct 51 tothe diffuser 52 (shown in FIG. 6a).

PRESSURE CONTAINING CASING ARRANGEMENT

Reference is again made to FIG. 4 and the enlarged gas-generator portiongiven in FIG. 5 showing the high-pressure casing arrangement of thisinvention. The gas generator casing 142 retains the compressor dischargepressure in cavity 144 which surrounds annular combustor liner 146. Asecond casing 148 surrounds gas-generator casing 142 wherein a dualannular cavity 150 and 151 is formed. Cooling steam fills cavity 150/151as will be subsequently explained. A third outer casing 152 surroundscasing 148 to form dual annular cavities 153 and 154 wherein bleed airfrom the axial compressor is deadended, as will be discussed.

A curved forward casing 156 attaches to the gas generator casingcircumferentially at gas-generator compressor flange 158 (See FIG. 3)forming a leak-proof joint to prevent steam leakage. Inner casing 148 isthin-walled and flexible to allow for differential thermal and pressureexpansion. Insulation 160 covers the entire outer surface. Fuel is fedto combustor 30 by fuel line 34.

Outer casing 152 is substantially thick enough to prevent pressuredistortion or blow-out with an adequate safety factor and is stiffenedby ribs 162 as required. Forward casing 156 likewise has a thick enoughsection 164 to prevent pressure distortion but as a thin enough sidewall 166 to allow for some flexibility. Forward casing 156 has three ormore annular grooves 168 wherein seal piston rings 170 fit similar tothose of a piston of a large reciprocating engine. The grooves 168 aredeep enough for piston rings 170 to compress and expand. Another set ofgrooves 168' and rings 170' are located at the rear of the gas generatorat connection ring 172. The outer casing 152 is thus allowed to floataxially and radially so to speak as the temperature of the gas generatorcasings 142 expands. This differential expansion can be considerable atthe temperature involved. Stiffening and support ring 174 is thin-walledand can have an expansion loop if required. It can also be noted thatorifices 176 allow deadended compressor air to flow back and forth inforward cavity 153 and rear cavity 154.

Casings 148, 152, 156 and associated parts can all be fabricated out ofhigh-strength steel and no expensive stainless or exotic materials arerequired such as may be required for the gas generator casing 142.

Piston rings 170 seal the outer casing 152 to the matching parts 164 and172 to prevent leakage but yet allow relative movement axially andradially between the parts. The rear diameter of casing 152 and flange172 and rings 170' can be slightly smaller than the diameter of theforward parts 152, 164 and 170 to allow cylindrical casing 152 to beslipped into place from the rear. Assembly and disassembly of the outercasing 152 is thus made quite easy and fast.

Locking pins 178 which are free to float secure and anchor the casing152 to part 164 and subsequently to gas generator compressor casingflange 158. There can be 2 or more of pins 178 equally spaced aroundcasing 156.

Piston-ring seals are used in gas turbines and other apparatus forexpansion joints and are not unique in themselves, but this applicationof the double piston sealing ring arrangement is quite different thanthose previously known or disclosed.

Alternately, referring to FIG. 6a, the thermal expansion between the gasgenerator casing 142 and the outer casing 152 can be accomodated bybellows-type expansion joints 180 and 182 with connecting flanges 184and 186 for assembly and disassembly.

The casing arrangement of this invention allows a substantiallyconventional gas generator with thin walls to be used for much higherpressures than normal. As an example, the LM5000 gas-generator casing isdesigned for 30 atmospheres of pressure. When operating at the higherpressure through intercooling and supercharging, the casing will operateat 40 to 60 atmospheres and will not be strong enough to preventdistortion and possible rupture without the strong outer casingarrangement.

Reference is again made to FIGS. 3 and 6a showing the compressor bleedair line 188 which pressurizes dual cavities 153 and 154. The airpressure will be somewhat higher than half that of the compressordischarge. As an example, for a 50 atmosphere discharge pressure, thecorrect compressor stage is selected to provide about a 30 atmosphericpressure to surround casing 148 and pressurize outer casing 152 throughcavities 153 and 154. The outer casing 152 thus is subjected to asubstantially lower pressure than the compressor discharge pressure.Inner casing 148 is subjected to the differential pressure of 50 minus30 or about 20 atmospheres. Steam chamber 150 is selected to be slightlymore than 50 atmospheres for the 50 cycle pressure ratio, as will bediscussed. The air pressure in cavities 153 and 154 has no outlet andflow is normally zero with the deadend situation. Flow only occursduring startup, load changes or shutdown to charge or discharge cavities153 and 154.

Cooling steam enters cavity 150 by means of flange connection 190 toring distribution header 192 then to connect to flexible tubing pipes194. The steam is produced by a heat recovery boiler and/or is extractedfrom a steam turbine at a pressure substantially equal to the compressordischarge pressure or preferably at a pressure 30 to 50 psi greater. Thesteam surrounds gas generator casing 142 and thus neutralizes theinternal-annular-discharge air pressure. In fact, the gas generatorcasing 142 can be placed in a slight condition of compression by the 30to 50 psi added pressure.

The steam is preferably dry and saturated for maximum cooling capacitywith maximum specific heat and the lowest temperature. However, thesteam can have 10° to 30° F. superheat to insure dryness as no waterparticles should be present at entrance through pipes 194. The steamcools the gas generator casing 142 as required. Baffles not shown can beapplied to control the flow around the casing 142 if required to controlcasing 142 cooling.

Again referring to FIG. 5, cooling steam in cavity 150 flows to innercasing 142 by means of open tubing 196. Cooling steam enters open tubing198 to cool the first-stage turbine disk 130 and blades 128. Coolingsteam enters nozzle vane opening 200, one or two for each vane, to cooleach first-stage nozzle vane 126. Cooling steam enters opening 202 tocontrol the cooling of gas generator casing rear section. Thus, cavity150 serves as an annular plenum or cooling steam distribution chamber.Hot steam can be taken out of turbine 20 by outlets 204, 206 and 208 foradmission back to the steam turbine, but the exit connections orexternal headers for these tubes are not shown. Alternately, this hotsteam can be piped by tubing not shown to rear steam cavity 151downstream of orifice 218. Cooling steam not used for internalgas-generator cooling is discharged at a higher temperature throughtubing 210 to ring header 212 and through flange 214.

Dam 216 with orifice or orifices 218, in number as needed, controldownstream steam pressure as required to lower the downstream pressureand control total steam flow to header 212.

Control of steam pressure and flow to cavity 150/151 during startup andshutdown of gas generator 20 can be by means of computer control of apressure regulating valve not shown that feeds steam to flange 190 andring header 192. Such control prevents over-pressurizing inner cavity150/151 and endangering blowout of inner casing 148 under transientconditions.

Reference is made to FIG. 6b showing a perspective view with partialsection of an intercooled reheat gas turbine incorporating a separateand independent industrial-type low-pressure compressor 24' coupled tothe high-pressure compressor 44 by coupling 220 and exhausting tointercoolers 38 and 40. Gas generator 20 exhausts through diffuser 52 toreheat combustor 50 and expands heated gas by power turbine 58 whichdrives load 60 by C shaft 62 being connected thereto by coupling flange59. The power turbine hot gas exhausts through exhaust hood 222 to aheat recovery boiler. Refer to FIG. 1. Diffuser 52 is connected to gasgenerator 20 by means of expansion joint 49 and is connected to reheatcombustor 50 by means of flange 53. Diffuser 52 can be a separateassembly which can be removed independently of gas generator 20 orreheat combustor 50 and power turbine 58 assembly.

Present-day aero-derivative gas generators such as the GELM500, GM 570K,R-R RB211 drive through the cold end, that is have an inner coaxialshaft extending through the gas generator with a coupling flangeconnection protruding out the center of the high-pressure air compressor44. Some heavy-duty non-aero industrial gas turbines likewise drive outthe cold (compressor) end. It is therefore possible to drive a separatelow-pressure compressor of this invention by such gas generators. Thisphysical arrangement, per se, is not unique. It is, however, unique todrive a specifically and specified low-ratio compressor of alow-pressure ratio of 1.8 to 3.2 for intercooling and specifically for acombined cycle for optimum combined-cycle efficiency. For instance, theJapanese Government reheat gas turbine AGT-J-100A intercools at a 4.85low-pressure ratio. Also, supercharging by large forced draft fans to 50inches of water and cooling by humidification has been used but not atthe specified low-pressure compressor ratio of this invention. Secondly,the pressurizing container required for aero-derivative light-weightcasings is likewise quite unique. No such process or arrangement hasever been proposed or contemplated before, to this inventor's knowledge.

CASCADE AIRFOIL DIFFUSER

Reference is now made to FIGS. 7a, 7b, and 8 showing a uniquelow-pressure compressor-diffuser arrangement of a cascade-airfoilconfiguration, which is part of this invention, to convert a greateramount of velocity head to static pressure than made possible by theapparatus of FIGS. 3 and 6a. As stated, the straight wall vanelessdiffuser of FIGS. 3 and 6b can only recover about 65 to 70 percent ofthe velocity head. The cascade airfoil diffuser of this invention canincrease this conversion from the 65 to 70 percent level toapproximately 85 to 90 percent. The process and associated apparatuswill now be explained. Convenient gas generator 20 duct connections canalso be provided, as will be explained, to install and remove said gasgenerator.

Air is discharged from the low-pressure compressor 24 through elbow 100.One or more turning vanes 224 are attached to elbow wall 226 by radialstruts 228. The turning vanes 224 prevent flow distortion andseparation. The radial discharge of the elbow 100 enters typically 12sections of essential rectangular cross-sectional ducts 230 each boundedby inner wall 232, outer wall 234 and side walls 236 and 238. There canbe fewer or more than 12 sections of duct 230 but typically there are12, 3 of which can be seen in FIG. 8. Side walls 236 and 238 can have anincluded divergent angle of zero to 30 degrees, but preferably no morethan 20 degrees, that is 10° on each side wall 236 and 238. Inner wall232 and outer wall 234 each curve outwardly like the outlet of a horn.The ratio of the rectangular cross-sectional area at the radial entrance240 to that of the exit 242 can be typically about 4 to effect avelocity change of 4. Typically the discharge velocity at exit 240 willbe about Mach 0.3 which translates to about 400 ft/sec for the flowingtemperature involved. The exit 242 velocity would be about 100 ft/sec(Mach 0.075).

Separation and thus channeling of the air causing eddy currents andvelocity-pressure-head dissipation is prevented by a series of two ormore stages of cascade airfoils. Three stages are shown in FIG. 7a. Thefirst stage has airfoils 246, the second stage has airfoils 248 and thethird stage has airfoils 250.

The airfoils are held in precise position by radial-duct sidewalls 236and 238. Welding can be used. Each inner duct is supported by struts252. The airfoils 246, 248 and 250 are precisely shaped according tospecific specifications provided by NASA or the RAF to produce thelowest drag and best lift (and thus diffusion) for the associatedvelocities. The surfaces of the airfoils are very smooth similar tostator vanes of an axial-flow compressor to reduce friction and loss.Referring to FIG. 7b, the airfoils are positioned to have a preciseangle of attack, β, at leading edge point A to provide the lift andassociated downstream diffusion with a low drag (pressure loss). Theairfoils diffuse the hot air as it flows from the maximum width point Band then to the trailing edge point C. Each stage of cascade airfoils isstaggered so that the wake of the upstream airfoils will not interferewith the downstream airfoil.

The hot air is diffused efficiently without separation by the cascadeairfoil arrangement. The air flow and direction thereof is controlled soit will fan out or spread apart The direction of flow of the air at theinner surface 234 is mostly tangential to the inner surface and the flowat the center pitch line is tangential to the center pitch line and theouter surface flow is tangential with the outer surface 234. Thedirection of the flow with respect to the radial position can be curvedbackwards to give space for flange 258, expansion joint 260 andtransition 262. The curvature can also be slightly forward with lesscurvature of elbow 100 and with the inlet 280 being moved fartherbackward to increase the curvature of the "S" shape of the inlet 280.

The gas, after leaving the cascade airfoil diffuser section, passesthrough an expansion joint 160 which has internal flow shielding notshown to form smooth internal surfaces. Flanges 264 and 266 connect theexpansion joint 260 to the ducting. The air then passes through thetransition piece 262 which changes the rectangular duct cross-section toa round cross-section at round flange connection 268. The air isdiffused further in the process and the exit velocity at flange 268 isnow about 50 to 75 ft/sec, or the desired velocity for low pressure lossflow for the piping and ducting 36 to the intercoolers 38 and 40 and thereturn ducting to the high-pressure compressor 44. The air, afterleaving transition piece 262, enters a toroidal-shaped ring-type duct270 with two outlet connections 272, one on each side of ring duct 270.The air flow is aided by internal turning vanes 274. Flanges 276 areconnected to line 36 which ducts the air to intercoolers 38 and 40.

The cooled air is returned to gas generator 20 by return air line 42which is connected to said gas generator by flange connection 278 one oneach side of said gas generator 20. The "S" walled inlet ducts 280 whichare on each side of said gas generator 20 have a generouscross-sectional area to provide a low velocity inlet of 50 to 75 ft/secor as required for low-pressure loss. Guide vanes 282 held by struts 284guide the air to the high-pressure compressor 44 as can be seen in FIG.7a.

The exit flange 276 and the inlet flange 278 can be readily disconnectedto allow gas generator 20 to be installed and removed for overhaul.Toroidal-shaped ring duct 270 need not be removed. Transition pieces 262and a section of return duct and associated expansion joint not shownthat connects line 42 to inlet flange 278 can be removed to allowclearance. FIG. 8 shows the transition ducts 262 as having been removed.

The straight-wall vaneless diffuser of FIGS. 3 and 6b will only providean effective area ratio change of about 2 without flow separation andthe exit velocity will be about 200 ft/sec resulting in a maximum ofsome 70 percent velocity pressure recovery when considering the velocityhead varies as the square of the velocity. Also the outlet and inletducts for the configuration shown in FIG. 3 extend radially outwardwhich necessitates a longer radial distance making it more clumsy todisconnect the gas generator 20 from outlet lines 36 and inlet lines 42.The design and assembly of the gas generator 20 becomes more complicatedand disassembly for overhauls would likewise be more difficult. Thereare these additional advantages favoring the apparatus arrangement shownin FIG. 7a, 7b and 8.

The gas generator 20 of this invention consists of individual modularsections which are assembled one to another to form a complete assemblywhich is essentially cylindrical in shape. The low-pressure compressor24 with discharge duct 36 or 230 module fits to the high-pressurecompressor 44 and return inlet duct 42 or 280 modules. The high-pressuresection 44/42 or 280 fits to the combustor section 30. The combustorsection 30 fits to the high-pressure turbine 46 assembly which then fitsto the low-pressure turbine assembly 28. The shafting and bearing parts26, 48, 80, 84 and 86 and their associated sub-parts are assembled withthe aforementioned modules to connect the rotating parts together. Theouter-casing assembly consisting of parts 48, 152, 164 and 166 andassociated sub-assembly parts fit over the high-pressure section of gasgenerator 20.

All these assemblies or sub-modules comprise one complete total modulethus forming gas generator 20.

Such a modular assembly of aero-derivative gas generators is well knownby those skilled in the art. Only the addition and arrangement of (a)the exit elbow 100, diffuser 102 or 230 and duct connection 36, (b) thereturn parts consisting of the duct connection 42, the "S" shaped duct104 or 280 turning vanes 282 and the like and (c) the pressurizingcasing parts 148, 152, 164 and 166 are new, all made possible by theoversized low-pressure compressor with a longer radius, r. This uniquearrangement allows more or less standard "core" parts ofhigh-temperature high-technology to be wedded with the lesser technologylow-pressure compressor, the duct connections and the pressurizingcannister assembly to form the complete cylindrically-shaped module ofgas generator 20.

THERMODYNAMIC ANALYSIS

The thermodynamic analysis of the heat-rate advantage of this inventionwill now be presented pinpointing the best intercoolinglow-pressure-compressor pressure-ratio range for optimum combined-cycleefficiency and the specific intercooling pressure for optimumcombined-cycle efficiency of this invention. Increased power outputderived by intercooling will be set forth.

The general reheat-gas-turbine cycle arrangement is given in theschematic diagram of FIG. 1. Air is compressed in a low-pressurecompressor 24 which is driven coaxially by turbine 28. Said air isdiffused and then ducted to intercoolers 38 and 40 where said air iscooled before being ducted back to be further compressed byhigh-pressure compressor 44 driven by turbine 46. High-pressure air at40 to 60 atmospheres is heated in first combustion chamber 30, is thenexpanded through turbines 46 and 28, is subsequently fully diffused bydiffuser 52 and then reheated in second combustion chamber 50 at a totalpressure of 4 to 9 atmospheres, depending on the cycle-pressure ratioand amount of steam-cooling/injection and is finally expanded toatmospheric pressure through power turbine 58 to drive load 60. The hotexhaust gases generate steam in a boiler for a conventional 2400 psigreheat-steam turbine not shown. The reheated

The intercoolers can be made in two sections 38 and 40 with condensatebeing used to cool the high end and cooling water used to cool the lowend as shown in FIG. 1. The following expected and typical pressurelosses are assumed for the intercooler. These losses are expressed aspercentages of the compressor intercooler total pressure:

    ______________________________________                                        LPC Diffuser Loss - %                                                                             .50                                                       LPC Exit Loss - %   .25                                                       IC Loss - %         2.00                                                      HPC Entrance Loss - %                                                                             .25                                                       Total Loss - %      3.00                                                      ______________________________________                                    

Incremental parasitic heat and mechanical losses that must be assignedto the incremental additional work saved by the intercooling process arealso associated with the intercooler. These losses are assumed to beconstant at five percent of the gross incremental work saved and arelisted as follows:

    ______________________________________                                        Generator Loss       Bearing Loss                                             Combustion Loss      Air Leakage                                              Radiation Loss       Auxiliary Loss.                                          ______________________________________                                    

The intercooler is sized to effect an exit temperature of 100° F. whenconsidering a standard 59° F. day. The minimum approach temperature isthus considered to be 41° F.

The adiabatic compression efficiencies of both the low-pressure andhigh-pressure compressors are assumed to be 88 percent. No attempt ismade to increase the efficiency at the low-pressure end or decrease theefficiency at the high end to simulate what actually takes place.

Careful studies of the non-intercooled reheat-gas-turbine combined cyclehave produced data for the three curves shown in FIG. 9. Note that threedifferent gas-turbine initial and reheat firing temperatures are shown,namely: (a) 2400°/2100° F., (b) 2600°/2200° F. and (c) 2600°/2400° F.Also, as can be seen, two different steamreheat temperature levels aregiven for the 2400 psig initial pressure. These are: 1000°/1000° F. and1100°/1050/850° F. for the double reheat. Higher steam and gastemperature can be considered.

Each curve representing the gas-turbine and steam-turbine-temperatureconditions peaks out at a specific cycle-pressure ratio as shown atpoint A 38 cycle pressure ratio (CPR), B 44 CPR and C 48 CPR. Thesethree curves are considered as the basic standards for comparing theintercooled cycle in terms of overall maximum-cycle efficiencyobtainable for the conditions given.

Referring to FIG. 1 again, it is assumed that the work saved bycompression intercooling is extracted through A shaft 26 and that therest of the reheat-gas-turbine cycle remains unchanged with theexception of incremental heat being added to the first combustor to heatthe air back to the original compressor-discharge temperature associatedwith the nonintercooled compression for any given cycle-pressure ratio.

This procedure greatly simplifies the analysis and neglects the smallvariations introduced by the heat required to vaporize and/or heat thefuel, the expansion work of the fuel itself and the slight increase inthe second combustor pressure due to the savings in compression work andthe incremental fuel-expansion work when considering a fixed exhausttemperature. Also, the very small amount of low-level heat recovery bythe condensate, if this scheme is used, is neglected for the sake ofsimplicity.

Finally, with reference to FIG. 9 the three basic non-intercooled gasturbines employ steam as the blade and combustor coolant and therefore,a lower compressor-discharge temperature resulting from the intercoolingand resulting cooling-air fluctuations does not enter into thereheat-gas-turbine cycle to distort the results.

There are three basic thermodynamic formulas to apply in calculatingcompression work and temperature rise. The first formula deals with thechange in enthalpy of the air and thus the work of compression: ##EQU1##where (H₂ -H₁) is the enthalpy change and work required, w is the weightflow, C_(p) is the specific heat considered to be constant, (T₂ -T₁) isthe change in temperature resulting from the compression at constantentropy and n is the adiabatic compression efficiency.

The second formula gives the relationship of temperature and pressureratio as follows: ##EQU2## where T₁ and T₂ are the absolute temperaturebefore and after compression, P₂ is the absolute pressure aftercompression, P₁ is the absolute initial pressure and k is the gasconstant (ratio of the two specific heats (C_(p) /C_(v)) This formula isbased on a constant entropy compression with no losses, thus, at 100percent efficiency.

The third formula is a combination of the two previous formulas andgives the temperature rise of compression which is directly related tothe work of compression and the heat content of the air: ##EQU3## Thespecific heat of air (C_(p)) is considered to have a constant value of0.24 BTU/lb - °F. (1.005 KJ/Kg - °C.) and the gas constant k isconsidered to be 1.4.

The aforestated three formulas and assumptions will produce closeresults which are valid for incremental compression work saved throughintercooling and incremental heat added after compression to raise thetemperature back to the nonintercooled level. More exact results usingthe gas tables are not warrented for comparison and relative compressionwork and changes in cycle efficiencies.

Formulas (2), (3) and (4) can be used to develop plots of temperatureversus pressure ratio as given in FIG. 10 where, for an inlettemperature of 59° F., the outlet temperature for any given ratio isplotted against the natural log of the pressure ratio. The natural logof pressure ratio is used to shrink the higher pressure-ratio scale andbroaden the lower ratio scale where needed for closer analysis. Theactual pressure ratios are given by the lower scale for easy reference.

Air is compressed from point 1 to point 5 without intercooling. Thislocus of temperature points is used as a standard for comparison.

Considering intercooling, air is first compressed from point 1 to point2 at which point the air is cooled by the intercooler to point 3. Theair is then compressed further to point 4. The total work of compressionis represented by the sum of the two temperature differences:accordingly (T₂ -T₁)+(T₄ -T₃). This temperature difference summation isgiven as point 6 in FIG. 10. Obviously the incremental work saved by theintercooling process is represented by (T₅ -T₆).

The extra heat required to heat the air back to temperature T₅ withoutintercooling(and thus the incremental heat to be added)is represented by(T₅ -T₄) in FIG. 10.

The intercooling pressure at point 2 can be varied from a pressure ratioof 1 to 60, for purposes of analysis, giving complete data ofincremental work saved and incremental heat added. The incremental cycleefficiency is then readily calculated using formula (5) whichincorporates the 5 percent parasitic losses: ##EQU4## where Ei is theincremental cycle efficiency of the intercooling process and where thenumerical subscripts refer to FIG. 10.

This analysis assumes that the compression work saved will behypothetically extracted out shaft A of FIG. 1 so that the rest of thecycle is not disrupted. This assumption is valid for purposes of ageneral analysis within the scope of accuracy of the other assumptions.The total cycle would have to be evaluated for precise accuracy wheremore exact values of compression efficiency, expansion efficiency,pressure losses and the like are known for a specific design.

First, considering no intercooler pressure loss and considering the airto be cooled all the way to the standard inlet temperature of 59° F., agraph of incremental cycle efficiency versus the natural log of cyclepressure ratio is obtained as given in the top lines C of FIGS. 11 and12 for two total cycle-pressure ratios of 40 and 60, the general area ofconcern. Note that FIG. 10 presents an example for a CPR beingintercooled at a 4 pressure ratio.

It can be seen that the incremental cycle efficiency is maximum at verylow intercooling pressures and falls off as the intercooling pressurerises. The efficiency is zero at full cycle pressure because noincremental work is saved. Also it can be noted as ascertained from FIG.9 that the combined cycle efficiency range is from 55 to 60 percent asshown by the shaded area of FIGS. 11 and 12. Therefore the only wayintercooling can possibly improve or equate to the known combined cycleefficiency is in the low intercooling pressure-ratio range of about 1.8to 3.2 when considering high-turbine inlet temperatures above 2400° F.Beyond a pressure ratio of 3, the combined cycles of FIG. 9 (points A, Band C) will be degraded in proportion to the amount of incremental worksaved and the incremental efficiency derived.

Using the model of FIG. 10 developed from Formula (5) and the zerointercooler pressure-loss target curves C of FIGS. 11 and 12, morerealistic plots of incremental efficiency versus intercooling pressureratio can now be developed for both the 40 CPR and 60 CPR cases. Asstated earlier, a 3 percent total intercooler pressure loss is believedto be realistic and is assumed. Also, the air is assumed to be cooled to100oF and alternately to 150° F. for comparison to the theoreticalcurves C of FIGS. 11 and 12. Reference is made to FIG. 11 for the 40 CPRcase and FIG. 12 for the 60 CPR case where an intercooling ratio span of1 to 10 is explored

A shaded area representing the combined-cycle efficiency range of 55 to60 percent taken from FIG. 9 can be seen at the top of each graph. Ascan be noted, the incremental cycle efficiency for intercooling (40 CPRand 60 CPR) for both the 100° F. and the 150° F. high-pressurecompressor (HPC) inlets peak out below the combined-cycle efficiencybeing considered as 55 percent efficiency, (points A and B) whichindicates that intercooling can only degrade combined-cycle efficiencyrelative to the amount of incremental compression work saved.

It can be noted that the incremental cycle efficiency plunges to zero atthe critical low-intercooling pressures because of the assumed pressureloss of 3 percent and the inlet temperature to the HPC exceeding thenormal compression temperature.

The most important consideration is at what intercooling pressure theincremental cycle efficiency peaks; that is points A and B of each ofFIGS. 11 and 12. The maximum efficiency as can be observed occurs atrather low return pressure, P₃, to the HPC and is about 1.8 for 100° F.and 40 CPR and about 2.0 for 100° F. and 60 CPR. The curves remainsomewhat flat and then fall off at a more constant rate as theintercooler pressure increases to 2.5 and 3 respectively.

A further observation made is that the intercooling incremental cycleefficiency lines for both 100° F. and 150° F. for the 3 percent pressureloss never cross the theoretical top lines C representing 59° F. HPCinlet and zero pressure drop. It can be concluded that minimum pressureloss and minimum HPC return temperature should be sought to yieldmaximum incremental efficiency.

There is another factor that must be considered to obtain the overallcombined-cycle degradation. This factor is the amount of incrementalcompression work saved by the intercooling process. Reference is made toFIG. 13 which is a plot of incremental work in BTU/lb of airflow versuspressure ratio for both the 40 CPR and 60 CPR cases and for 100° F. HPCinlet and 3 percent pressure loss.

It can be seen that maximum work saved by intercooling occurs at thesquare root of the total pressure ratio, points A and B of FIG. 13. Theincremental specific work, W_(i), saved can be determined by applyingformula (6) following: ##EQU5## where C_(p) is the specific heatconsidered to be 24 BTU/lb - °F. (1.005 KJ/Kg - °C.) for standard airand where the numerical subscripts refer to FIG. 10.

An intercooler pressure range of 1.8 to 3.2 is shown in the range areawhere incremental cycle efficiency remains at a relatively high level.These lower intercooler pressures must be considered for maximumcombined-cycle efficiency, but nevertheless at some sacrifice in overallcombined cycle efficiency and moreover at less than maximum incrementalgas turbine and combined-cycle output. The ordinate has an additionalscale for percent gas turbine power increase and another for percentcombined-cycle power increase based on calculated values of 303.36BTU/lb net work for the gas turbine and 121.49 BTU/lb net work for thesteam turbine.

Intercooling does increase gas turbine output from about 6 to 14 percentand combined cycle output by 4 to 10 percent for the range shown in FIG.13.

The effect intercooling has on combined cycle efficiency can now bedetermined by applying the known calculated data of a combined cycle forboth the 40 CPR and 60 CPR cases for 100° F. HPC inlet tabulated below:

    ______________________________________                                                          BTU/lb                                                      ______________________________________                                        Gas Turbine Output  303.36                                                    Steam Turbine Output                                                                              121.49                                                    Total Fuel Input (LHV)                                                                            740.03                                                    Net Cycle Efficiency (LHV)     57.41%.                                        ______________________________________                                    

The incremental values for intercooling are added to the above valuesand then equated to percentages. The results are given in FIG. 14 wherecombined-cycle efficiency loss is plotted against the intercoolerpressure ratio (outlet of the low-pressure compressor).

The range to be considered is shown to be about 1.8 to 3.2 as shown inFIG. 14. The combined cycle degradation can be as great as 4.5 percentat a six pressure ratio, but for practical considerations for a morereasonable intercooler pressure ratio of about 2.5 for 40 CPR thedegradation would be about 1.2 percent and about 1.0 percent for the 60CPR at a pressure ratio of about 3.

These values can now be applied to FIG. 9 to arrive at projected overallcombined-cycle efficiency. As an example, considering a 50 CPRintercooler pressure of 2.5 and conditions of line B of FIG. 9 the netcombined-cycle efficiency would be as follows:((57.3 (1-0.008)=56.8percent LHV)) with reference to point X of FIG. 14. If the intercoolingpressure ratio is raised to 4 then the combined cycle efficiency wouldfall thus: ((57.3(1-0.0250)=55.9 percent LHV)) with reference to point Yof FIG. 14.

The above examples show how important intercooling pressure ratio iswith regard to combined cycle efficiency. It should, however, be pointedout that at the 4 intercooler pressure ratio the combined-cycle outputis projected to be increased by about 12 percent for a given airflow inaccordance with FIG. 13 whereas the increase in output at the lower 2.5pressure ratio would only be about 8 percent.

There is one final additional consideration regarding intercooling whichdeals with the gas-generator combustion temperature rise (GGTR) and theHPC discharge temperature (CDT). FIG. 15 is a plot of the GGTR and theCDT versus the intercooler pressure ratio for both the 40 CPR and the 60CPR cases.

As the intercooler pressure ratio rises, the GGTR also rises as shownand conversely the CDT falls. More fuel is burned in the gas-generatorcombustor as the intercooler pressure increases.

The intercooler pressure ratio range to be considered for bestcombined-cycle efficiency is given in FIG. 15 and corresponding GGTRsand CDTs can be compared in regard to changes in combustion radiationheat, combustor cooling and NO_(x) control. It is worthy to note that asthe GGTR becomes greater with more fuel being burned, the oxygen contentof the gas to the reheat combustor will decrease and the gas will alsobe vitiated with steam from blade cooling and NO_(x) control. Therefore,the safe flamability limits and the complete combustion could become aproblem at some point.

Intercooling according to the teachings of this invention degradescombined-cycle efficiency. This degradation as given in FIG. 14 startsfrom about 0.8 percent, point X at the optimum intercooling ratio rangeof 1.8 to 2.2 and increases rapidly to about 2.5 percent at a 4 ratiopoint Y for the 41° F. interrapidly cooled approach temperature to theatmospheric temperature (100°-59°=41° F.). Cooling to a closertemperature would reduce this loss only slightly.

The term effectiveness is often used to describe heat transferparameters and surface areas. Effectiveness is the ratio of thetemperature reduction achieved divided by the full potential to anambient or coolant datum. Some studies have indicated an effectivenessas high as 95 percent which would yield a very low approach temperatureof some 10° F. instead of the 41° F. for the 100° F. inlet to the highpressure compressor. Of course a lower return temperature would reducethe degradation according to the teaching of this invention and inparticular with reference to FIGS. 11 and 12. However, the lower theapproach temperature and the higher the effectiveness, the higher willbe the intercooler pressure loss, offsetting the efficiency gain;greater intercooler heat transfer surface area will be required. The 41°F. approach seems reasonable because there is little to be gained ingoing to a lower approach temperature in terms of gain in combined-cycleefficiency.

FIGS. 11, 12 and 14 clearly show that the best intercooler pressureratio for maximum efficiency is very close to 2.0 and when consideringextra power saved according to FIG. 13 this ratio can be extended toabout 2.5 without excessive combined-cycle efficiency loss. This ratiocan be further extended to 3.2 before cycle efficiency degradationbecomes excessive.

The gas generator of this invention has the capital cost advantages offactory pre-assembly and test. One single module assembly is madepossible whereby the low-pressure compressor and high-pressurecompressor with the coaxial shafting can be made up of variousindividual modules of an aero-derivative approach. There is no flexiblecoupling between the two compressor components and only one bearing isneeded for the low pressure compressor. Radial and longitudinal space ismade available for air exit and readmittance because of the largerradius of the low-pressure compressor.

The aero-derivative high-pressure casings with the advantages of thinparts for low thermal stresses can be retained through the adaptation ofthe heavier external pressurizing casing. The thick casing andassociated parts can be fabricated out of ordinary high-strength steeland will not be expensive. Manufacturing costs savings can be realizedby using more or less standard designs of existing and future fanengines. Interchangeability of parts is another advantage of the gasgenerator of this invention.

The pressurizing container placed over the high-pressure gas generatorthin casing makes it possible to apply existing aero-derivative gasgenerators that can drive an industrial-type low-pressure compressorthrough the cold end of the gas generator thereby broadening the scopeof this invention. The gas generator can be adapted to burn both liquidand gaseous fuels and there is a particular advantage in considering agaseous fuel derived from coal. Very high combined-cycle efficiencies of55 to 60 percent are calculated for the complete combined cycleincorporating the gas generator of this invention.

The thermodynamic analysis shows that the low-pressure compressorpressure ratio must be low and about 1.8 to 3.2 to prevent excessivecombined cycle degradation when the gas generator of this inventionfurnishes hot gas under pressure to a reheat combustor and then to apower turbine.

The gas generator and complete reheat gas turbine of this invention canbe applied to generator electric power or produce mechanical power todrive a large ship where efficiency and space are important factors tobe considered. The quick change-out feature of the gas generator makessuch applications attractive.

According to the teachings of this invention, the particular existingaero-derivative gas generators cited and the increase in "core" flowmade possible by the intercooling and supercharging effect, the reheatgas turbines of this invention can have an air-flow rate ranging from 50lb/sec to as high as 400 lb/sec. However, the air flow rate andresulting output can be higher and extend upwards to 500 lb/sec or more,only limited to design bottlenecks encountered such as the size of thelast-stage blade of the power turbine and the cooling thereof.

The air-flow rates given above will produce a reheat gas turbine outputranging from 15,000 KW to 150,000 KW for existing aero gas generatorscited when adapted for intercooling and supercharging. Future unitscould yield even greater power output. The larger sizes of 100,000 to200,000 KW are the desireable sizes for future electric utilityinstallations as the present plans are to increase existing plant outputby increments of 100,000 to 200,000 KW and not going to the largersingle units of the past decades of 1,000 MW size. It should beremembered that the reheat gas turbine arrangements of this inventionwill produce about one-third more secondary power from the steam orvapor bottoming portion of the cycle to supplement existingelectric-utility power plant output.

While the invention has been described in connection with specificembodiments thereof, it will be understood that it is capable of furthermodifications, and this application is intended to cover any variations,uses or adaptations of the invention following, in general, theprinciples of the invention and including such departures from thepresent disclosure as come within known or customary practice in the artto which the invention pertains and as may be applied to the essentialfeatures hereinbefore set forth, and as fall within the scope of theinvention and the limits of the appended claims.

That which is claimed is:
 1. In a power producing system comprising atwin spool gas generator and a power turbine, said gas generator havinga low pressure compressor driven by a low pressure turbine, a highpressure compressor driven by a high pressure turbine, a combustorpositioned between said high pressure compressor and said high pressureturbine, said power turbine positioned downstream from said low pressureturbine, the improvement being characterized in that: said high and lowpressure turbines being axially positioned and independently rotatablefor driving said high and low pressure compressors respectively by meansof concentric coaxial outer and inner shafting respectively said gasgenerator including at least one externally mounted intercoolerpositioned between said low pressure compressor and said high pressurecompressor, at least one compressor outlet duct from said low pressurecompressor communicating with said intercooler and at least one returnduct from said intercooler communicating with said high pressurecompressor, wherein said compressor outlet and return ducts andconnections between said compressors and said intercooler are providedbetween said axially positioned low and high pressure compressors forair flow to and from said intercooler in counterflow with coolant, saidoutlet duct being configured to radially expand said air flow to a lowvelocity and said return duct being configured for low radial flowreturn velocity to said high pressure compressor.
 2. The power producingsystem as defined in claim 1 including a heat recovery boiler means anda steam turbine or vapor turbine means to form a combined cycle, whereinsaid power turbine exhausts into said heat recovery boiler meansproducing steam for expanding in said steam turbine or vapor turbinemeans.
 3. In a power producing system comprising a twin spool gasgenerator, a reheat combustor and a power turbine, said gas generatorhaving a low pressure compressor driven by a low pressure turbine, ahigh pressure compressor driven by a high pressure turbine and a firstcombustor positioned between said high pressure compressor and said highpressure turbine, said reheat combustor being positioned between saidlow pressure turbine and said power turbine, the improvement beingcharacterized in that: said high and low pressure turbines being axiallypositioned and independently rotatable for driving said high and lowpressure compressors respectively by means of concentric coaxial outerand inner shafting respectively, said gas generator including at leastone externally mounted intercooler positioned between said low pressurecompressor and said high pressure compressor, at least one compressoroutlet duct from said low pressure compressor communicating with saidintercooler and at least one return duct from said intercoolercommunicating with said high pressure compressor, wherein saidcompressor outlet and return ducts and connections between saidcompressors and said intercooler are provided between said axiallypositioned low and high pressure compressors for air flow to and fromsaid intercooler in counterflow with coolant, said outlet duct beingconfigured to radially expand said air flow to a low velocity and saidreturn duct being configured for low radial flow return velocity to saidhigh pressure compressor.
 4. The power producing system as defined inclaim 3 including a heat recovery boiler means and a steam turbine orvapor turbine means to form a combined cycle, wherein said power turbineexhausts into said heat recovery boiler means producing steam forexpanding in said steam turbine or vapor turbine means.
 5. The powerproducing system as defined in claim 4 including a condenser means forcondensing steam from said heat recovery boiler means.
 6. The powerproducing system as defined in claim 5 including means in saidintercooler communicating with said condenser means whereby the coolantfor said intercooler is said condensate from said condenser means.
 7. Ina power producing system comprising a twin spool gas generator, a powerturbine, a heat recovery boiler means, and a steam or vapor expansionturbine means, said gas generator having a low pressure compressordriven by a low pressure turbine, a high pressure compressor driven by ahigh pressure turbine and a combustor positioned between said highpressure compressor and said high pressure turbine, said power turbinepositioned downstream from said low pressure turbine, said heat recoveryboiler means positioned downstream from said power turbine, theimprovement being characterized in that: said high and low pressureturbines being axially positioned and independently rotatable fordriving said high and low pressure compressors respectively by means ofconcentric coaxial outer and inner shafting respectively, said gasgenerator including at least one intercooler means positioned betweensaid low pressure compressor and said high pressure compressor, at leastone compressor outlet duct from said low pressure compressorcommunicating with said intercooler means and at least one return ductfrom said intercooler means communicating with said high pressurecompressor, wherein said compressor outlet and return ducts andconnections between said compressors and said intercooler means areprovided between said axially positioned low and high pressurecompressors for said flow to and from said intercooler means, whereinsaid low pressure compressor is an axial flow design staged to produce apressure ratio no greater than about 3 and said high pressure compressoris an axial flow design of 11 or more stages being fed with cooler andmore dense air by said intercooler means to produce a discharge pressureof at least about 40 atmospheres.
 8. In a power producing systemcomprising a twin spool gas generator, a reheat combustor, a powerturbine, a heat recovery boiler means and a steam or vapor expansionturbine means, said gas generator having a low pressure compressordriven by a low pressure turbine, a high pressure compressor driven by ahigh pressure turbine, and a first combustor positioned between saidhigh pressure compressor and said high pressure turbine, said reheatcombustor being positioned between said low pressure turbine and saidpower turbine, said heat recovery boiler means positioned downstreamfrom said power turbine, the improvement being characterized in that:said high and low pressure turbines being axially positioned andindependently rotatable for driving said high and low pressurecompressors respectively by means of concentric coaxial outer and innershafting respectively, said gas generator including at least oneintercooler means positioned between said low pressure compressor andsaid high pressure compressor, at least one compressor outlet duct fromsaid low pressure compressor communicating with said intercooler meansand at least one return duct from said intercooler means communicatingwith said high pressure compressor wherein said compressor connectionsbetween said compressor and said intercooler means are provided betweensaid axially positioned low and high pressure compressors for air flowto and from said intercooler means, wherein said low pressure compressoris an axial flow design staged to produce a pressure ratio no greaterthan about 3 and said high pressure compressor is an axial flow designof 11 or more stages being fed with cooler and more dense air by saidintercooler means to produce a discharge pressure of at least about 40atmospheres.